1. Field of the Invention
The present invention relates generally to fuel control method and system for a cylinder injection type (or direct injection type) internal combustion engine for a motor vehicle in which fuel is injected directly into engine cylinders. More particularly, the present invention is concerned with fuel control method and system for the cylinder injection type internal combustion engine equipped with an exhaust gas recirculation system (EGR system for short) in which a fuel injection control mode (hereinafter also referred to simply as the injection mode) can be changed over from a mode in which fuel injection is performed during a compression stroke of the engine for realizing a high air-fuel ratio (hereinafter referred to as the compression-stroke fuel injection mode) to a mode in which fuel injection is performed during a suction stroke of the engine for realizing a low air-fuel ratio (hereinafter referred to as the suction-stroke fuel injection mode) while ensuring improved combustion performance of the engine softwarewise by resorting to correspondingly improved control procedure without need for additional provision of especial hardware component(s) to this end.
2. Description of Related Art
For having better understanding of the principle underlying the invention, technical background thereof will be described in some detail. FIG. 4 is a schematic diagram showing a conventional fuel control system for an internal combustion engine in which a fuel injector is installed within an intake passageway or manifold (mounted outside of the cylinder).
Referring to FIG. 4, an internal combustion engine (hereinafter also referred to simply as the engine) for a motor vehicle is generally denoted by a reference numeral 1.
The engine 1 is equipped with an intake passageway in which an intake air flow sensor 2 is installed at a position upstream of the engine for measuring the air flow (an amount of intake air) Qa fed to the engine 1.
Further mounted within the intake passageway is a throttle valve 3 which is operatively coupled to an accelerator pedal (not shown) adapted to be manipulated by a driver of the motor vehicle and which serves for regulating the air flow supplied to the engine 1 in dependence on the depression stroke of the accelerator pedal.
For the purpose of detecting the angular position of the throttle valve 3 as a throttle opening degree .alpha., a throttle opening degree sensor 4 (which may also be referred to as the throttle valve position sensor) is provided in association with the throttle valve 3.
Provided in association with a crank shaft of the engine 1 is a crank angle sensor 5 for detecting a rotation speed (rpm) of the engine 1 as well as a rotational position or angular position of the crank shaft. Thus, the crank angle sensor 5 generates a crank angle signal SGT as the output signal thereof from which information concerning the rotation speed (rpm) of the engine 1 as well as the information concerning the angular position of the crank shaft can be derived.
Temperature Tw of cooling water for the engine 1 is detected by a water temperature sensor 6 which thus can serve as a means for detecting warmed-up state of the engine 1.
An O2-sensor 7 is provided in association with an exhaust passageway of the engine 1 for detecting an oxygen concentration or content Do (corresponding to the air-fuel ratio) of the exhaust gas discharged from the engine 1.
An electronic control unit (also referred to as ECU for short) 8 is provided for arithmetically determining various control quantities in dependence on the operation states of the engine 1. The ECU 8 constitutes a major part of an engine control system and serves for deciding the operating states of the engine 1 on the basis of detection signals outputted from the various sensors installed in the engine at relevant locations. The ECU unit 8 is designed to generate a control signal in dependence on the operation state of the engine for realizing combustion of air-fuel mixture at a desired air-fuel ratio.
A spark plug 9 is installed within the combustion chamber of each of the engine cylinders and undergoes conventional ignition control.
As can be seen in FIG. 4, provided in parallel to the intake passageway across the throttle valve 3 is a bypass passage which an air bypass valve 10 is installed for selectively opening or closing the bypass passage. Thus, it is possible to control the engine rotation number or engine speed (rpm) even when the throttle valve 3 is fully closed (i.e., even when the engine operates in the idling mode).
Further, the air bypass valve 10 can be used for the engine torque control in the running state of the motor vehicle.
A fuel injector 11 is mounted within the intake passageway at a position upstream of the engine 1 for injecting the fuel within the intake passageway.
Further, in order to control the amount of the exhaust gas to be recirculated (also referred to as the exhaust gas recirculation or EGR quantity) into the combustion chamber of the engine 1 with a view to reducing the content of nitrogen oxides or NOx carried by the exhaust gas, there is provided an EGR control system a part of which is constituted by an EGR valve 12.
A cylinder identifying sensor 13 is provided in association with a cam shaft interlinked to the crank shaft to serve for outputting a cylinder identifying signal SGC for identifying discriminatively the cylinder within which the combustion takes place.
The detection signals Qa, .alpha., SGT, Tw, Do and SGC derived from the outputs of the intake air flow sensor 2, the throttle opening degree sensor 4, the crank angle sensor 5, the water temperature sensor 6, the O2-sensor 7 and the cylinder identifying sensor 13, respectively, are inputted to the ECU unit 8 as the information indicative of the operating state of the engine 1.
On the other hand, the various components such as the spark plug 9, the air bypass valve 10, the fuel injector 11 and the EGR valve 12 are driven in response to control signals P, B, J and G, respectively, which represent correspondingly the control quantities determined arithmetically by the ECU unit 8.
As can be seen in FIG. 4, in the fuel control system of the conventional internal combustion engines known heretofore, the fuel injector 11 is mounted within the intake passageway of the engine 1. In recent years, however, the cylinder injection type fuel control system (also known as the direct injection type fuel control system) which allows the fuel to be directly injected into the engine cylinder has been developed.
Since the cylinder injection type fuel control system can promise advantageous and profitable effects such as mentioned below, the cylinder injection type fuel control system is very attractive as the ideal fuel injection control system for the engine of a motor vehicle.
(1) Reduction of the content of harmful gases contained in the exhaust gas
In general, in the indirect fuel injection type internal combustion engine in which the fuel is injected externally of the cylinder, a part of fuel as injected is likely to be deposited on the intake valve and wall of the intake passageway. For this reason, it is necessary to take into consideration the amount of fuel likely to be deposited before charging the fuel into the cylinder, particularly when the engine is started from a low-temperature state where the fuel is difficult to vaporize or when the engine is in a transient operation mode in which fuel supply has to be changed at a relatively high rate. By contrast, in the case of the cylinder injection type internal combustion engine, the air-fuel ratio can be increased so that the air-fuel mixture becomes lean without taking into consideration the delay involved in the transportation of the fuel, whereby contents of harmful HC (hydro carbon) gas and CO (carbon monoxide) gas carried by the exhaust gas can be reduced.
(2) Reduction of fuel cost
In the cylinder injection type internal combustion engine, the fuel is injected immediately before the ignition timing, whereby there is formed a mass of combustible fuel mixture around the spark plug 9 at the time of ignition. In other words, the gas mixture containing the fuel is distributed nonuniformly. Thus, the fuel-air mixture undergoes a so-called stratified combustion. Consequently, the air-fuel ratio in appearance between the amount of air and that of the fuel charged into the engine cylinder can be significantly increased, which means that the air-fuel mixture can be made remarkably lean.
Besides, owing to the realization of the stratified combustion, combustion of the air-fuel mixture is scarcely affected adversely even when the exhaust gas is recirculated with an increased ratio. By virtue of this feature and additionally the reduction of pumping loss, the fuel-cost performance of the engine can be enhanced significantly.
(3) Realization of high output power of the engine
Owing to the stratified combustion of the air-fuel mixture concentrated around the spark plug 9, the amount of end gas (i.e., the air-fuel mixture gas in the regions located remotely from the spark plug 9) decreases, which is favorable to the improvement of the anti-knocking performance of the engine. Thus, the compression ratio of the internal combustion engine can be increased.
Furthermore, because the fuel is converted into gas or gasified within the cylinder, the intake air is deprived of heat as vaporization heat. Consequently, the density of the intake air can be increased with the volumetric efficiency being enhanced, which in turn promises high output of the engine.
(4) Enhancement of drivability
By virtue of the direct fuel injection into the cylinder in the cylinder injection type engine system, the time lag intervening between the fuel injection and the generation of output torque by the engine is short when compared with the cylinder injection type engine system.
Thus, there can be realized the internal combustion engine system which is capable of responding speedily to the demand of the driver.
Now, description will be made of a conventional cylinder injection type internal combustion engine system for having better understanding of the invention. FIG. 5 is a schematic diagram showing generally a structure of a conventional fuel control system for a cylinder injection type internal combustion engine such as described in Japanese Unexamined Patent Application Publication No. 187819/1992 (JP-A-4-187819). In the figure, components like as or equivalent to those described hereinbefore by reference to FIG. 4 are designated by like reference characters and repeated description in detail of these components is omitted.
In the cylinder injection type internal combustion engine system, measures for improving the combustion performance of the internal combustion engine is incorporated in the engine itself.
Referring to FIG. 5, a fuel injector 11 is mounted within a cylinder of the engine 1 at high pressure side.
A fuel injector driver 14 is interposed between the ECU 8 and the fuel injector 11 to drive the fuel injector 11 at high speed and high pressure in response to a control signal J issued by the ECU unit 8.
At this juncture, comparison of the cylinder injection type internal combustion engine shown in FIG. 5 with that described hereinbefore by reference to FIG. 4 shows that the former differs structurally from the latter in that the fuel injector 11 for supplying the fuel is not mounted within the intake manifold but installed within the cylinder (i.e., in the combustion chamber) of the engine 1.
Parenthetically, in the case of the fuel control system for the cylinder injection type engine, the fuel injector 11 is implemented with high-speed/high-pressure specifications so that the fuel can be injected into the cylinder with high pressure within a short preceding time period in the suction stroke and the compression stroke. Thus, the fuel control system shown in FIG. 5 also differs from the system shown in FIG. 4 in that the injector driver 14 for driving the fuel injector 11 is additionally provided.
Next, operation of the conventional fuel control system for the cylinder injection type internal combustion engine shown in FIG. 5 will be elucidated by reference to a timing chart shown in FIG. 6 and a flow chart illustrated in FIG. 7 together with FIG. 8 showing relevant control data structure.
FIG. 6 is a timing chart for illustrating changeover of fuel injection mode M between a compression-stroke injection mode MA (i.e., mode in which the fuel is injected directly in the cylinder during the compression stroke) and a suction-stroke injection mode MB (i.e., mode is which the fuel is injected during suction stroke) together with changes in the air-fuel ratio A/F, fuel injection timing Tj of the fuel injector 11, ignition timing Tp of the spark plug 9, EGR quantity Qg and the intake air flow or quantity Qa, respectively, which take place upon changeover of the fuel injection mode M. In FIG. 6, the compression-stroke injection mode MA is validated for the combustion of excessively lean air-fuel mixture while the suction-stroke injection mode MB is validated for the combustion of enriched air-fuel mixture.
The fuel injection mode M is changed over from the compression-stroke injection mode MA to the suction-stroke injection mode MB at a time point t1. Further, at a time point t2, the fuel injection timing Tj and the ignition timing Tp are changed over to the timings for the suction stroke injection mode to the timings for the compression-stroke injection mode. The air-fuel ratio A/F becomes stable at a time point t3. The air-fuel ratio in the compression-stroke injection mode MA is represented by A/FA with the air-fuel ratio in the suction-stroke injection mode MB being represented by A/FB. Further, A/Fr shown in FIG. 6 represents a reference value for the air-fuel ratio on the basis of which the time point t2 is determined. Furthermore, reference symbol QgA represents the EGR quantity in the compression-stroke injection mode MA, QgB represents the EGR quantity in the suction-stroke injection mode MB, QaA represents the intake air flow or quantity in the compression-stroke injection mode MA, and QaB represents the intake air quantity in the suction-stroke injection mode MB.
As can be seen from FIG. 6, the fuel control system for the cylinder injection type internal combustion engine has two fuel injection modes, i.e., the compression-stroke injection mode MA and the suction-stroke injection mode MB.
In the compression-stroke injection mode MA, the fuel is supplied to the engine 1 during the compression stroke to effectuate the stratified combustion in the over-lean state (i.e., the state in which the air-fuel mixture is excessively lean) in order to enhance emission and fuel consumption features of the engine. On the other hand, in the suction-stroke injection mode MB, the fuel is injected during the suction stroke. In that case, the ordinary combustion of uniform mixture is realized, whereby the engine output power can be increased.
FIG. 7 is a flow chart for illustrating operation sequence or control program stored in a microcomputer or microprocessor incorporated in the ECU unit 8 shown in FIG. 5. Further, FIG. 8 shows desired values of various control quantities in the form of two-dimensional map as a function of the engine rotation number Ne (rpm) and the engine load Le. For instance, there are shown map data values of the desired air-fuel ratio A/Fo, desired fuel injection timing Tjo, desired ignition timing Tpo, desired EGR quantity Qgo and the desired intake air quantity Qao, respectively.
Referring to FIG. 7, the ECU unit 8 makes decision concerning the fuel injection mode M of the engine 1 on the basis of the information of the intake air quantity Qa, the throttle opening degree .alpha., the crank angle signal SGT, the cooling water temperature Tw, the oxygen concentration Do of the exhaust gas and the cylinder identifying signal SGC outputted, respectively, from the relevant sensors installed on the engine 1 (step S1), to thereby determine whether or not the current fuel injection mode M is the suction-stroke injection mode MB (step S2).
When it is decided in the step S2 that the engine 1 is in the suction-stroke injection mode MB (i.e., when the decision step S2 results in affirmation "YES"), the ECU 8 arithmetically determines or calculates the control quantities, i.e., the desired air-fuel ratio A/Fo, the desired fuel injection timing Tjo, the desired ignition timing Tpo, the desired EGR quantity Qgo and the desired intake air quantity Qao for the suction-stroke fuel injection (refer to FIG. 8) in a step S3.
At this juncture, it should be mentioned that the desired values mentioned above are previously calculated as the values set separately for the suction-stroke fuel injection and the compression-stroke fuel injection, respectively, in dependence on the engine rotation number Ne (rotation speed in rpm) and the engine load Le (represented ordinarily by the intake air quantity Qa in each combustion cycle).
Subsequently, tailing processing for decreasing gradually the desired air-fuel ratio A/Fo is executed to enrich the air-fuel mixture (step S4), which is then followed by a step S5 where it is decided in the step S5 whether or not the desired air-fuel ratio A/Fo is greater than a reference value A/Fr (i.e., whether or not the air-fuel mixture is lean, to say in another way).
When it is decided that the desired air-fuel ratio A/Fo is greater than the reference value A/Fr (i.e., in case the decision step S5 results in "YES"), the calculated values of the desired fuel injection timing Tjo and the desired ignition timing Tpo for the compression-stroke fuel injection are adopted (step S6), whereon the procedure exits from the processing routine shown in FIG. 7 and proceeds to a succeeding processing.
On the other hand, when it is decided in the step S5 that the desired air-fuel ratio A/Fo is equal to or smaller than the reference value A/Fr (i.e., in case the decision step S5 results in negation "NO"), the step S6 is skipped, and the values of the desired fuel injection timing Tjo and the desired ignition timing Tpo calculated for the suction-stroke fuel injection in the step S3 are adopted, whereon the procedure leaves the processing routine illustrated in FIG. 7 to proceed to the succeeding processing.
By contrast, when it is decided in the aforementioned step S2 that the engine 1 is not in the suction-stroke injection mode MB but in the compression-stroke injection mode MA (i.e., when the decision step S2 results in negation "NO"), the desired air-fuel ratio A/Fo, the desired fuel injection timing Tjo, the desired ignition timing Tpo, the desired EGR quantity Qgo and the desired intake air quantity Qao for the compression-stroke fuel injection are affirmatively determined as the control quantities on the basis of the engine rotation number (engine speed) Ne and the engine load Le, as mentioned previously in conjunction with the step S3, whereon the control procedure leaves the processing routine shown in FIG. 7 and proceeds to a succeeding processing.
Through the procedure comprised of the processing steps S1 to S7 described above, the tailing processing (i.e., gradually decreasing processing) of the air-fuel ratio A/F is started at the time point tl at which the fuel injection mode M is changed over to the suction-stroke injection mode MB from the compression-stroke injection mode MA, as shown in FIG. 6.
Further, at the time point t2 at which the air-fuel ratio A/F becomes equal to or smaller than the reference value A/Fr, the fuel injection timing Tj and the ignition timing Tp are changed over from the timings for the compression-stroke combustion to the timings for the suction-stroke combustion.
In conjunction with the changeover of the fuel injection mode M from the compression-stroke injection mode MA to the suction-stroke injection mode MB, it is however noted that the combustion is likely to be unstable because not only the air-fuel ratio A/F and the fuel injection timing Tj but also the control quantities such as the ignition timing Tp, the EGR quantity Qg and the intake air quantity Qa are varied.
By way of example, when the fuel injection mode M is changed over from the compression-stroke injection mode MA to the suction-stroke injection mode MB, as shown in FIG. 6, then, the air-fuel ratio A/F, the EGR quantity Qg and the intake air quantity Qa are varied from the control quantities A/FA, QgA and QaA for the compression-stroke fuel injection to the control quantities A/FB, QgB and QaB for the suction-stroke fuel injection, whereon the succeeding control is started.
In that case, in connection with the control of the air-fuel ratio A/F, the tailing operation from the air-fuel ratio A/FA to the air-fuel ratio A/FB is started in order to suppress abrupt or shock-like variation of the engine torque.
On the other hand, the fuel injection timing Tj and the ignition timing Tp can instantaneously be changed over to the fuel injection timing Tj and the ignition timing Tp for the suction stroke combustion from to those for the compression stroke combustion at the time point t2 at which the air-fuel ratio A/F becomes equal to or smaller than the predetermined reference value A/Fr (i.e., at the timing point when the air-fuel mixture becomes lean) because there are two stable time points in respect to the combustion in the compression stroke and the suction stroke.
Furthermore, in conjunction with the control of the EGR quantity Qg and the intake air quantity Qa, it is noted that upon changing-over of the fuel injection mode M from the compression-stroke fuel injection mode to the suction-stroke fuel injection mode, some time lag is involved for the EGR quantity QgA and the intake air quantity QaA to reach the desired EGR quantity QgB and the desired intake air quantity QaB (desired control quantities) after the changeover of the fuel injection mode M, respectively.
However, since the intake air quantity Qa supplied to the engine 1 is measured by means of the intake air flow sensor 2 (see FIG. 5), the lag involved in feeding the intake air to the engine upon the mode changeover mentioned above can be measured. Accordingly, the lag in feeding the intake air does not exert any appreciable influence to the control of the air-fuel ratio A/F because the latter is controlled on the basis of the measured value derived from the output of the intake air flow sensor 2.
By contrast, the EGR quantity Qg which is so set that the combustion can take place without fail can vary as the intake air quantity Qa and the air-fuel ratio A/F change upon changeover of the fuel injection mode M. In other words, the EGR quantity Qg changes in dependence on both the air-fuel ratio A/F and the intake air quantity Qa which vary upon every changeover of the fuel injection mode M.
Consequently, there may arise such situation in which the combustion becomes unstable although it depends on the combination of the variable parameters such as the air-fuel ratio A/F, the intake air quantity Qa and others.
Additionally, it is noted that in the fuel control system for the cylinder injection type internal combustion engine, the fuel is supplied to the engine cylinder immediately before the ignition timing so that the stratified combustion can occur in the compression-stroke injection mode MA as described hereinbefore. Accordingly, the air-fuel ratio of the air-fuel mixture around the spark plug 9 in the actual combustion is close to the stoichiometric ratio of 14.7, even though the air-fuel ratio of the mixture as supplied is thirty or more.
Certainly in the indirect injection type fuel control system in which the fuel is injected into the intake passageway (see FIG. 4), combustion takes place with the air-fuel ratio of ca. 20 (lean burn) after the intake air and the fuel have been mixed uniformly. By contrast, in the fuel control system for the cylinder injection type internal combustion engine, combustion in the compression-stroke injection mode is performed by firing the air-fuel mixture existing around the spark plug 9 and having the air-fuel ratio of ca. 16. Consequently, in the fuel control system for the cylinder injection type internal combustion engine, greater amount of nitrogen oxides (NOx) is discharged when compared with the indirect injection type engine. Such being the circumstances, in the cylinder injection type internal combustion engine, a large amount or quantity of exhaust gas is recirculated to the engine with a view to realizing reduction of nitrogen oxides (NOx).
As will now be apparent from the foregoing description, in the fuel control system for the cylinder injection type internal combustion engine known heretofore such as disclosed in Japanese Unexamined Patent Application Publication No. 187819/1992 (JP-A-4-187819), combustion in the compression-stroke injection mode MA is carried out through combination of the stratified combustion which can be realized by subtle control of the fuel injection timing Tj and the ignition timing Tp and recirculation of a large amount or quantity of exhaust gas which may lead to degradation of the combustion in the ordinary indirect injection type engine 1 (see FIG. 4).
On the other hand, in the suction-stroke injection mode MB which can ensure high output torque of the engine, the fuel injection is performed similarly to that of the indirect injection type engine for thereby effectuating uniform mixture combustion.
When the fuel injection mode M is changed over to the rich combustion state in the suction-stroke injection mode MB from the over-lean combustion state in the compression-stroke injection mode MA, the air-fuel ratio A/F, the EGR quantity Qg, the fuel injection timing Tj and the ignition timing Tp are correspondingly changed under the control of the ECU and at the same time the quantity of intake air fed to the engine 1 is decreased by controlling correspondingly the air bypass valve 10 (see FIG. 5) in order to prevent the output torque of the engine from fluctuation which may otherwise be brought about by the changeover from the compression-stroke injection mode MA to the suction-stroke injection mode MB.
In this manner, for changing over the fuel injection mode M to the suction-stroke injection mode MB from the compression-stroke injection mode MA, a large number of control parameters have to be changed simultaneously in order to afford the change or variation of the combustion state.
In conjunction with the simultaneous changeover of many control parameters, it is however noted that because of nonuniformity in the performance among the components subjected to the control with the respective parameters, age changing thereof, variation in the environmental conditions during the running of the motor vehicle and/or difference of the combustion states, there may arise such situation in which the combustion state can not make transition smoothly to the combustion of uniform mixture from the stratified combustion, incurring possibly unstable combustion and hence fluctuation of the rotation speed of the engine 1 upon changeover of the fuel injection mode M.
Under the circumstances, such changeover control is performed for the control parameters that upon changeover of the fuel injection mode M of the engine 1, parameters such as the EGR quantity Qg and the intake air quantity Qa the control of which is usually accompanied with lag are controlled in precedence to the change-over control of the fuel injection timing Tj and the ignition timing Tp so that the changeover controls of the individual components as involved can be performed effectively at a same time.
In the fuel control system for the cylinder injection type internal combustion engine among others, recirculation of a large quantity of exhaust gas is carried out in many operation ranges of the engine. Accordingly, even when the EGR control is performed effectively simultaneously with the control of the other relevant components upon changeover of the fuel injection mode M, there may actually occur such situation that the air-fuel mixture undergoes combustion in the course of change of the EGR quantity Qg (and hence with various EGR quantities), presenting a problem that unstable state of combustion may be brought about.
As will now be understood from the above, in the fuel control system for the cylinder injection type internal combustion engine, the changeover control of the fuel injection mode is performed such that the fuel injection timing Tj, the ignition timing Tp and other parameters are changed under control after the EGR quantity Qg and the intake air quantity Qa have been controlled in precedence.
However, the EGR quantity Qg introduced into the engine may vary continuously even after the changeover control of the fuel injection timing Tj and the ignition timing Tp, giving rise to a problem that the unstable combustion state may thereby be incurred.